1. Field of the Invention
The present invention relates to a hydrostatic bearing, and in particular to a variable capillary apparatus for a hydrostatic bearing and a motion error compensating method using the same which can control a pocket pressure of a hydrostatic bearing, thereby compensating motion errors of a hydrostatic table caused by geometrical errors of a variety of precise measurement units and precise machine tools which are incorporated with the hydrostatic table.
2. Description of the Prior Art
A hydrostatic bearing is a non-contact type bearing that guides fluids such as air, a lubrication oil or the like to an inside thereof and obtains a load capacity by a restricting effect. The hydrostatic bearing attenuates an erroneous low frequency vibration transmitted along a spindle shaft or guide rails by an averaging effect of a fluid film formed between a confronting member and itself, and also attenuates a high frequency vibration by its high damping characteristics. So, the hydrostatic bearing is adequate to the spindle shaft and a transporting table of the precise machine tools so that it is utilized in the spindle shaft and guideway of various grinding machines and in the guideway of precise machine tools. FIG. 1 shows an operating principle of a hydrostatic bearing 100 and an equivalent electric circuit thereof. A table at which bearing 100 is installed linearly moves along a guideway 120. The table maintains a predetermined interval from guideway 120 by a pressurized fluid film injected from a pocket 130 formed at an upper surface thereof. A plurality of pockets can be formed at the transporting table of the machine tools. The fluid flow into the pockets is regulated such that the fluid can apply a proper pressure in response to a load applied to the table. In case of a single pad type hydrostatic bearing, bearing gap h is evaluated from the principle that the sum of a load W applied to the table including the weight of the table is equal to the integration of pressure P formed between the guideway and the table with respect to an entire bearing area A, as shown in equation (1). That is, ##EQU1##
In equation (1), A, B are a nondimentional area and a flowrate defined by the pocket area with respect to the entire bearing area; Ps is a supply pressure of a fluid supply device; Pr is a pocket pressure; and kc is a capillary coefficient of the capillary device for giving a stiffness to the hydrostatic bearing.
Thus, A, B increase with the pocket area.
Although the hydrostatic bearing has advantages such as low friction and a high moving accuracy due to an averaging effect, it has a disadvantage such as low stiffness. To improve the stiffness of the hydrostatic bearing, it is effective to minimize the bearing gap. In addition, to optimally design a bearing with a small bearing gap, the amount of a fluid that flows in the bearing should be restricted corresponding to the bearing gap.
However, when the bearing gap decreases, the machining accuracy of the bearing should be improved. Thus, there is a limitation on the improvement of the stiffness of the bearing by means of a decrease of the bearing gap. To overcome the limitation, various riable restricting mechanisms have been proposed.
As an example of such mechanisms, a diaphragm type variable restricting bearing 200 or 250 as shown in FIG. 2 or 3 is known. In this example, bearing surfaces 210 and 260 are composed of elastically deformabe diaphragms. The shapes of bearing surfaces 210 and 260 are deformed corresponding to pressure changes on bearing surfaces 210 and 260 due to load changes applied to the bearings 200 and 250. Thus, a high bearing stiffness can be accomplished. Particularly, in the mechanism shown in FIG. 3, the inner periphery of the diaphragm 260 is elastically supported by an O-ring 270. The pressure of the rear surface of the diaphragm 260 varies corresponding to the pressure on the bearing surface 260 through a small hole 290 formed nearly at the center of the housing. Thus, the shape of the diaphragm sensitively varies corresponding to the variation of the load. Such a diaphragm type variable restricting bearing has been disclosed heretofore.
When the above-described diaphragm type variable restricting mechanism is used with proper bearing conditions, high bearing stiffness can be accomplished. However, the above-described diaphragm type variable restricting mechanism has the following disadvantages: (a) since the bearing surface should be deformed, the applications of the bearing are limited; (b) when the diaphragm is largely deformed, a pocket is formed in the bearing. Thus, the diaphragm may self-oscillate due to a compression of air; and (c) it is difficult to machine the diaphragm. Depending on the machining accuracy, the bearing characteristics vary, thereby affecting the durability of the diaphragm.
Meanwhile, in manufacturing the spindle shaft and the table using the hydrostatic bearing, further in the case of the spindle shaft, a manufacturing precision must be achieved in its circular and cylindrical contours. And in the case of the table, the manufacturing precision must be achieved in a straightness, a flatness, and a right angled degree of each rail, and a degree of a parallel layout and a degree of a perpendicular layout in view of a couple of parallel rails. So, although we can expect the averaging effect of the fluid film, it is difficult to achieve a displacement precision having an order of a micron. In addition, since a motion error due to the manufacturing error depends on the geometrical shape of the table and rails, even if the peripheral devices have high precision, it is difficult to compensate for the motion error.
In the prior art, three types of compensation have been proposed. Fixed compensation uses a capillary or orifice to act as a fixed-value resistance. Variable compensation includes the use of diaphragms and valves to provide a flow inversely proportional to the pocket resistance, thereby creating a larger pressure differential than created with the use of a fixed compensation device. Both of these types of compensation, however, must be tuned to the bearing gap.
As smaller and smaller bearing gaps are sought in order to increase the performance of the bearing, manufacturing errors make the use of either of these types of compensation more and more difficult by requiring hand-tuning of the compensation device. Since a machine tool with three axes may have 36 bearing pockets, the labor required becomes prohibitive.
A third type of compensation is called self-compensation because it uses the change in bearing gap to allow the bearing to change the flow of fluid to the pockets by itself. Existing self-compensation methods have utilized linear passageways on the face of the bearing and have been directed primarily to applications in spindles. These design have not, however, proven themselves to provide acceptable performance in the commercial section because of inefficient flow patterns that are difficult analytically to determine, particularly the flow field near the end of the linear grooves, often resulting in improper resistance design and which then require hand-tuning of the compensator. Difficulty has also been experienced with prior linear groove self-compensation units because the geometry has not always been realistically implementable.